This type of bearing essentially is comprised of three specific types of zones constituting its geometry. The first zone would be the entrance portion to the orifice. The second zone is the restriction, which can be either a very small orifice in which the fluid control is by its diameter, or a capillary (small tube in which the fluid control is by its diameter and length). The third zone is the exit portion from the orifice. In a plain type bearing, this zone does not exist. In most bearing designs, this zone consists of a very thin disc shape.
The manner in which hydrostatic bearings work is that as the shaft approaches the orifices, because it is somewhat restricted the flow, there will be a pressure increase between the orifice area and shaft. Likewise, 180 degrees from the just mentioned orifice, the shaft will move away from the corresponding orifice. This has the effect of allowing the fluid to flow more freely causing a decrease in pressure between the orifice and the said shaft. Thus a restoring force is created which tends to keep the shaft centered in its bearing. The minimum number of orifices required per circumference would obviously be three. Normally four to 32 orifices are utilized. Another consideration concerning the design of these hydrostatic bearing systems is pneumatic hammer or “water hammer” if a fluid is used. This phenomena is basically a vibration which is set up between the shaft and its bearing. One of the main causes of this unwanted vibration is due to the time necessary for the so-called fluid compensated bearing to restore the shaft to its initial position. The fluid in association with the orifices and spacing caused by the shaft movement, sets up the occurring restoring forces. However, there is a finite amount of time required for the fluid to reach an equilibrium depending on the shaft position in its bearing. This so-called “time constant” is what causes the abovementioned vibration instabilities. In order to prevent this, very small orifices are necessary, along with the corresponding geometry such as the previously mentioned exit zone.
In general, gas lubricated bearings are normally assumed to have a given spring constant and a small amount of inherent dampening due to viscous forces. This spring constant may be measured by plotting a force vs. displacement curve. The damping coefficient may be determined by ascertaining the decrease in amplitude as a function of time of the bearing system after it is subjected to an impact load.
However, gas bearings are different from more conventional bearings in the sense that there is a finite time lag between the initial application of a displacement force and the time required for the bearing to reach a steady state condition, i.e., new position. Upon extensive investigation of experimental evidence, it was noted that this time constant or time lag is difficult to measure because it is obscured by the bearings damping. This time constant can be calculated, and becomes larger if the pocket type bearing is not designed with this in mind. It can be shown by computer simulation that, in general, the longer the time constant (first-order lag), the more unstable the system becomes. This time lag is caused by the fact that a finite amount of time is required for the fluid to flow out of the bearing or into the bearing depending on the location of the shaft until a new pressure distribution has been reached. It is known that a pocket type bearing has a greater restoring force, i.e., the force tending to establish a balancing equilibrium condition.